Deswirl mechanisms and roller bearings in an axial thrust equalization mechanism for liquid cryogenic turbomachinery

ABSTRACT

Vane, fin, and hole arrangements establish a predetermined reduced swirl at the inlet of mechanical seals and the inlet of a variable axial orifice gap which act in harmony as an axial thrust equalizing system for use in liquid cryogenic turbines and pumps. In said establishment the stiffness, damping, and inertia in said seal in conjunction with said variable orifice gap is manipulated, including the destabilizing cross-coupled stiffness which is reduced. Said seal is of either labyrinth annular type formed by a plurality of teeth, annular smooth, or a plurality diamond annular surface pattern. Said variable orifice gap is smooth. Liquid for the axial thrust equalizing seal is initially bled from the main to pass through a preset deswirl mechanism. The deswirl mechanism consists of either a plurality of vanes, fins, grooves, or circular holes that guide liquid radial inward before passing through said mechanical seal. After exiting the seal said liquid passes through a second deswirl mechanism consisting of a plurality of vanes, fins, or grooves before entering a variable axial orifice gap. The variable orifice moves in axial position to variably restrict balancing liquid and generate backpressure in the pressure chamber to balance the axial thrust caused by a plurality of impellers on the same single shaft. After passing through the variable orifice the bleed liquid can pass past a sealed lubricated roller bearing for heat exchange to cool said bearing with the cryogenic liquid along grooves in a bearing liner. Alternatively the liquid can also pass directly through an open unsealed bearing for cooling.

RELATED APPLICATIONS

This Application is related to U.S. Provisional Patent Application Ser.No. 60/920,618 filed Mar. 29, 2007 entitled DESWIRL MECHANICS AND ROLLERBEARINGS IN AN AXIAL THRUST EQUALIZATION MECHANISM FOR LIQUID CRYOGENICTURBOMACHINERY, which is incorporated herein by reference in itsentirety, and claims any and all benefits to which it is entitledtherefrom.

FIELD OF THE INVENTION

The present invention relates to liquid cryogenic centrifugal pumps andturbines of the submerged motor or generator type.

BACKGROUND OF THE INVENTION

Vertical cryogenic submerged motor pumps and submerged generatorturbines operate in the liquefied cryogenic gases industry. They aremost prominent in the liquid hydrocarbon industry for liquefied naturalgas, liquefied ethane gas, and liquefied propane gas. U.S. Pat. No.5,659,205 to Weisser, which is hereby incorporated by reference in itsentirety herein, teaches that due to the low cryogenic temperatures thisstyle of pump and turbine operates with the axial thrust of the rotatingassembly totally equalized to zero. U.S. Pat. No. 6,441,508 to Hylton isalso hereby incorporated by reference in its entirety herein.

To achieve this, a conventional axial thrust equalizing mechanism suchas shown in FIG. 1 is applied that uses pressurized bleed liquid whichis passed through a seal restriction at the back of the highest pressureimpeller. Afterwards this bleed liquid passes into a pressure chamberwhose pressure is controlled by a downstream variable area orifice. Thisorifice is variable in the axial direction along the pump and turbineshaft and comes from the rotating assembly which is designed to floataxially a small distance. In a condition of up axial thrust on therotating assembly, the variable orifice is pushed smaller. As such, thebleed liquid flow rate is reduced and the pressure drop across the bleedwear ring is reduced. The pressure in the downstream pressure chamberrises which increases the force on the impeller and rotating assembly sothat a reaction force is established to push the variable orifice largerand equalize the thrust. Contrarily, in a condition of down axial thruston the rotating assembly, the variable orifice is pulled larger. Assuch, the bleed liquid flow rate is increased and the pressure dropacross the bleed wear ring is increased. The pressure in the downstreampressure chamber decreases which decreases the force on the impeller androtating assembly so that a reaction force is established to push thevariable orifice smaller and equalize the thrust. So a net zero axialthrust is always established by the thrust equalization mechanism.

SUMMARY AND ADVANTAGES OF THE PRESENT INVENTION

The power rating of liquid cryogenic pumps and turbines in high pressureapplications continues to grow as motivated by customer demands. Thistranslates to higher power concentration machinery. So the axial thrustmechanism must balance larger thrust levels. Greater radial thrustlevels are also experienced which the seals must react to avoid overlylarger shaft deflections and overly large shaft diameters to compensate.Thus, means are sought to increase the stiffness of impeller flowinduced reaction forces to stiffen the shaft. Increasing the shaftdamping is also beneficial. Benckert, H., et al. teach in “Flow InducedSpring Coefficients of Labyrinth Seals for Application in RotorDynamics” published 1980, which is hereby incorporated by reference inits entirety, that means are also sought to reduce the well documenteddestabilizing cross-coupled stiffness in the mechanical seals. Overallincreasing the stiffness and damping while decreasing the cross-coupledstiffness will reduce rotordynamic whirl and vibrations. This can beseen from first principles with the equation of motion applied to arotating assembly experiencing small displacements δ in the x and ydirection written as follows:

${- \begin{bmatrix}{F_{x}(t)} \\{F_{y}(t)}\end{bmatrix}} = {{\begin{bmatrix}{k_{xx}(\delta)} & {k_{xy}(\delta)} \\{k_{yx}(\delta)} & {k_{yy}(\delta)}\end{bmatrix}\left\lbrack \begin{matrix}x \\y\end{matrix} \right\rbrack} + {\begin{bmatrix}{c_{xx}(\delta)} & {c_{xy}(\delta)} \\{c_{yx}(\delta)} & {c_{yy}(\delta)}\end{bmatrix}\left\lbrack \frac{\frac{x}{t}}{\frac{y}{t}} \right\rbrack} + {\quad{\left\lbrack \begin{matrix}{m_{xx}(\delta)} & {m_{xy}(\delta)} \\{m_{{yx}\;}(\delta)} & {m_{yy}(\delta)}\end{matrix} \right\rbrack \left\lbrack \frac{\frac{^{2}x}{t^{2}}}{\frac{^{2}y}{t^{2}}} \right\rbrack}}}$

Note: Both the direct coupled and cross coupled terms are represented inthe stiffness (k), damping (c), and inertia mass (m) matrix. For smalldisplacements, the coefficients in these equations are taken as linear.Separating the forcing contributions in the absolute reference frameresults in the following:

$\begin{bmatrix}{F_{x}(t)} \\{F_{y}(t)}\end{bmatrix} = {\begin{bmatrix}F_{xa} \\F_{ya}\end{bmatrix} + \begin{bmatrix}{F_{x}(t)} \\{F_{y}(t)}\end{bmatrix}_{whirl} + \begin{bmatrix}{F_{x}(t)} \\{F_{y}(t)}\end{bmatrix}_{nonwhirl}}$

The force contributions are dividing into steady and unsteady. Theunsteady force contribution is further subdivided into whirl and nonewhirl portions. The whirl contribution will be taken as a circular orbitthat experiences small periodic displacements of δ in x and y soδ=δ_(o)+iy and δ=δ_(o) exp(iω_(w)t). In this relation, ω_(w) is theimpeller whirl frequency. Now, expanding the previous equation for thewhirl terms gives the following:

$\begin{bmatrix}{F_{x}(t)} \\{F_{y}(t)}\end{bmatrix}_{whirl} = {\quad{\begin{bmatrix}\left( {{m_{xx}\omega_{w}^{2}} - {c_{xy}\omega_{w}} - k_{xx}} \right) & \left( {{m_{xy}\omega_{w}^{2}} + {c_{xx}\omega_{w}} - k_{xy}} \right) \\\left( {{m_{yx}\omega_{w}^{2}} - {c_{yy}\omega_{w}} - k_{yx}} \right) & \left( {{m_{yy}\omega_{w}^{2}} + {c_{yx}\omega_{w}} - k_{yy}} \right)\end{bmatrix}\begin{bmatrix}{\delta_{o}\cos \; \omega_{w}t} \\{\delta_{o}\sin \; \omega_{w}t}\end{bmatrix}}}$

It is now more convenient and intuitive to write this equation indimensionless form as follows:

$\begin{bmatrix}{F_{x}^{*}(t)} \\{F_{y}^{*}(t)}\end{bmatrix} = {\quad{\begin{bmatrix}\left( {{m_{xx}^{*}\frac{\omega_{w}^{2}}{\omega^{2}}} - {c_{xy}^{*}\frac{\omega_{w}}{\omega}} - k_{xx}^{*}} \right) & \left( {{m_{xy}^{*}\frac{\omega_{w}^{2}}{\omega^{2}}} + {c_{xx}^{*}\frac{\omega_{w}}{\omega}} - k_{xy}^{*}} \right) \\\left( {{m_{yx}^{*}\frac{\omega_{w}^{2}}{\omega^{2}}} - {c_{yy}^{*}\frac{\omega_{w}}{\omega}} - k_{yx}^{*}} \right) & \left( {{m_{yy}^{*}\frac{\omega_{w}^{2}}{\omega^{2}}} + {c_{yx}^{*}\frac{\omega_{w}}{\omega}} - k_{yy}^{*}} \right)\end{bmatrix}\begin{bmatrix}{\delta_{o}\cos \; \omega_{w}{t/R_{2}}} \\{\delta_{o}\sin \; \omega_{w}{t/R_{2}}}\end{bmatrix}}}$

The * designates use of the dimensionless quantities with F*=F/πρR₂³B₂ω², x*=x/R2, dx*/dt=(dx/dt)/R₂ω, and dx²*/dt=(d²x/dt²)/R₂ω². Thedimensionless stiffness, damping and added mass coefficients used arek*_(ij)=k_(ij)/πρR₂ ²B₂ω², c*_(ij)=c_(ij)/πρR₂ ²B₂ω, m*_(ij)=m_(ij)/πρR₂²B₂. This expression gives the x, y component of the forces but thegreater interest for turbomachinery vibrations lies in the tangentialand radial forces from the rotating assembly center. So we convert topolar coordinates with F*_(r)+iF*_(θ)=(F*_(x)+iF*_(y))exp(−iω_(w)t) andget the following equation:

$\begin{bmatrix}F_{\theta}^{*} \\F_{r}^{*}\end{bmatrix}_{whirl} = \begin{bmatrix}{{- {m_{xy}^{*}\left( \frac{\omega_{w}}{\omega} \right)}^{2}} - {c_{xx}^{*}\left( \frac{\omega_{w}}{\omega} \right)} + k_{xy}^{*} + {m_{yz}^{*}\left( \frac{\omega_{w}}{\omega} \right)}^{2} - {c_{yy}^{*}\left( \frac{\omega_{w}}{\omega} \right)} - k_{yx}^{*}} \\{{m_{xx}^{*}\left( \frac{\omega_{w}}{\omega} \right)}^{2} - {c_{xy}^{*}\left( \frac{\omega_{w}}{\omega} \right)} - k_{xx}^{*} + {m_{yy}^{*}\left( \frac{\omega_{w}}{\omega} \right)}^{2} + {c_{yx}^{*}\left( \frac{\omega_{w}}{\omega} \right)} - k_{yy}^{*}}\end{bmatrix}$

Now the rotation of the coefficients about the x, y axis is taken withisometry, which most whirl related test data supports, meaning the termswith subscript xx equal the yy terms and the subscript xy terms equalthe negative yx terms. This then gives the following equation:

$\begin{bmatrix}F_{\theta}^{*} \\F_{r}^{*}\end{bmatrix}_{whirl} = {2\begin{bmatrix}{{- {m_{xy}^{*}\left( \frac{\omega_{w}}{\omega} \right)}^{2}} - {c_{xx}^{*}\left( \frac{\omega_{w}}{\omega} \right)} + k_{xy}^{*}} \\{{+ {m_{xx}^{*}\left( \frac{\omega_{w}}{\omega} \right)}^{2}} - {c_{xy}^{*}\left( \frac{\omega_{w}}{\omega} \right)} - k_{xx}^{*}}\end{bmatrix}}$

For the circumferential force if F*_(θ) is negative, in the reversedirection of the impeller whirl rotation, an impeller whirl stabilizingforce is experienced. If F*_(θ) is positive, in the direction of whirl,this destabilizes the impeller by eliciting greater whirl. The stabilityboundary is found by taking the value of F*_(θ)=0 and m_(xy) asnegligible in the previous equation to give ω_(w)/ω=k*_(xy)/c*_(xx) asthe whirl ratio limit. Taking m_(xy) as negligible with respect to thestiffness and damping is reasonable for most but not all rotordynamicproblems, although it does illustrate the origins of the whirl ratiolimit. In dimensional form, this tangential whirl ratio limit as astability condition then simplifies to the following:

$\left( \frac{\omega_{w}}{\omega} \right)_{\theta \; {limit}} = \frac{k_{xy}}{c_{xx}\omega}$

Therefore, the tangential stability whirl ratio limit is a balancebetween cross coupled stiffness forces k_(xy) that drive the whirl anddamping forces c_(xx)ω that oppose the whirl. For a constant angularfrequency with a whirl larger than (ω_(w)/ω)_(θ limit), the tangentialforce acts in a stabilizing manner. For a constant angular frequencywith a whirl smaller than (ω_(w)/ω)_(θ limit) the tangential force actsin a destabilizing manner, unless the whirl orbit is backwards in whichcase this is stabilizing. Hence the desire to decrease the cross-coupledstiffness k_(xy) (and increase the direct damping) is beneficial forimproved whirl stability and reduced rotordynamic vibrations. Applyingthis finding, several research institutions and patents such as U.S.Pat. No. 5,190,440 to Maier have applied swirl brakes to labyrinth sealsin high temperature gas compressors.

It is this premise applied in conjunction with a thrust equalizationmechanism that is unique for liquid cryogenic pumps and turbines. In sodoing the benefit of a reduced destabilizing cross-coupled stiffness inthe seal and balance mechanism is gained. Further, the direct coupledstiffness k_(xx) is increased in the seal along with an increase in thedirect coupled c_(xx) damping. The reduced swirl in the variable orificeof the balance mechanism also provides an improved equalization of theaxial thrust with unwanted flow separation regions avoided in theorifice gap. So several advancements in thrust balancing devices forliquid cryogenic pumps and turbines are addressed with the claims ofthis patent.

Accordingly, there are provided herein several unique improvements tothe axial thrust equalizing mechanism which address the deficiencies ofpreswirl in the prior art of submerged motor liquid cryogenic pumps andturbines. The invention reduces the destabilizing cross-coupledstiffness while concurrently increasing the direct coupled stiffness anddirect coupled damping in the mechanical seals. This is achieved withinthe framework of an improved axial thrust equalization. The sealsthemselves consist of either labyrinth type, smooth type, or surfacepattern type such as diamond surface mesh. Holes are also claimed tolocally inject fluid with zero swirl and stop any residual swirlingliquid seal flow.

Another embodiment provides deswirl fins, vanes or grooves upstream ofthe variable orifice used for the thrust equalization. These ensure thevariable orifice receives liquid with adjusted prespecified preswirlwhich may be zero with the flow directed primarily radially. This avoidsfluid instabilities including separation near the orifice which cansuddenly collapse or form giving the thrust balance system a rapidchange in balance position. The predominately radial flow liquiddirection also improves the capacity of balancing higher thrust levelsneeded for more powerful pumps and turbines.

A further embodiment provides both a sealed and unsealed roller bearingoperating in conjunction with the axial thrust equalizing mechanism andthe deswirl devices. Currently unsealed roller bearings are the priorart. Sealed bearings packed with lubricants are not used in cryogenicapplications for fear of freezing. Recent advances in synthetic greasenow make available unfrozen grease down to temperatures of −60° C. Thisis applicable to liquid propane and butane pumps and turbines,particularly in situations where the fluid is dirty and can causereduced bearing life for an unsealed bearing. For situations where thefluid temperature is lower, a bearing heater and sensor are embodiedwhich briefly preheat the frozen grease before start-up. After start-up,the bearing heater may no longer be needed as the bearing itself maygenerate sufficient heat.

A last embodiment provides deswirl vanes, fins, or holes on the seals onthe plurality of impeller eyes and interstages. These are also useful toreduce the cross-coupled stiffness while concurrently increasing thedirect coupled stiffness and direct coupled damping. Surface patternssuch as diamond mesh are also utilized with a smooth rotating surfaceand a inlet deswirl mechanism for the same rotordynamic benefit.

Numerous other advantages and features of the present invention willbecome readily apparent from the following detailed description of theinvention and the embodiments thereof, from the claims and from theaccompanying drawings.

Benefits and features of the invention are made more apparent with thefollowing detailed description of a presently preferred embodimentthereof in connection with the accompanying drawings, wherein likereference numerals are applied to like elements.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a partial, cross-sectional overall view of a multistagecryogenic pump at the axial thrust equalizing mechanism highest pressureimpeller including seal deswirl mechanisms and lubricated sealed rollerbearing which are constructed in accordance with the invention.

FIG. 2 is a partial, cross-sectional view of four embodiments of thevaned or grooved or finned deswirl mechanism upstream of an impellerbackside labyrinth seal and the variable axial orifice gap constructedin accordance with the invention.

FIG. 3 is a partial, cross-sectional view of one embodiment of thedeswirl mechanism incorporating holes upstream of an impeller backsidelabyrinth seal with local injection into the seal constructed inaccordance with the invention.

FIG. 4 is a partial, cross-sectional view of one embodiment of thedeswirl mechanism incorporating a diamond surface pattern on the sealstator and smooth surface on the impeller backside constructed inaccordance with the invention.

FIG. 5 is a partial, cross-sectional view of four embodiments of thedeswirl mechanism operating in conjunction with the variable axialorifice gap using a cooled lubricated scaled roller bearing constructedin accordance with the invention.

FIG. 6 is a partial, axial view of two embodiments of the sealed rollerbearing liner with a bearing heater and temperature sensor for coldliquid applications constructed in accordance with the invention.

FIG. 7 is a partial, cross-sectional view of two embodiments of thedeswirl mechanism operating in conjunction with the variable axialorifice gap using a cooled dry lubricated unsealed roller bearingconstructed in accordance with the invention.

FIG. 8 is a partial, cross-sectional view of three embodiments with finsor vanes or grooves upstream of the impeller eye seal as a deswirlmechanism constructed in accordance with the invention.

FIG. 9 is a partial, cross-sectional view of one embodiment with adiamond surface pattern on the stator seal upstream of the impeller eyeseal as a deswirl mechanism constructed in accordance with theinvention.

FIG. 10 is a partial, cross-sectional view of three embodiments withfins or vanes or grooves upstream of the impeller interstage seal as adeswirl mechanism constructed in accordance with the invention.

FIG. 11 is a partial, cross-sectional view of one embodiment with adiamond surface pattern on die seal stator upstream of the impellerinterstage seal as a deswirl mechanism constructed in accordance withthe invention.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

The description that follows is presented to enable one skilled in theart to make and use the present invention, and is provided in thecontext of a particular application and its requirements. Variousmodifications to the disclosed embodiments will be apparent to thoseskilled in the art, and the general principals discussed below may beapplied to other embodiments and applications without departing from thescope and spirit of the invention. Therefore, the invention is notintended to be limited to the embodiments disclosed, but the inventionis to be given the largest possible scope which is consistent with theprincipals and features described herein.

It will be understood that while numerous preferred embodiments of thepresent invention are presented herein, numerous of the individualelements and functional aspects of the embodiments are similar.Therefore, it will be understood that structural elements of thenumerous apparatus disclosed herein having similar or identical functionmay have like reference numerals associated therewith.

Referring to FIG. 1 in particular, shown therein is a centrifugalimpeller apparatus of a centrifugal pump or turbine used for handlingcryogenic liquids. The apparatus incorporates annular labyrinth sealelements that are constructed in accordance with the invention. Forclarity, the various components of the centrifugal pump or turbineimpeller apparatus, as well as the annular labyrinth seal elements, areshown with sectional views of an upward portion thereof, it beingrealized that such elements are symmetrically oriented entirely aroundthe rotatable shaft center.

The case of a centrifugal pump is described realizing the reverse flowequivalent nature of a centrifugal turbine for which the claims alsohold. It will be understood that for purposes of the currentapplication, an LNG pump may be used to increase the pressure of theliquid LNG, while a turbine may act to lower the pressure of the liquidLNG. While the terms “pump” and “turbine” may be used interchangeably incertain portions of the current application, in general the primarydifferences between the two are described as follows: In the case of anLNG pump used to increase the pressure of the liquid LNG, flow of themain stream of liquid LNG will be into the pump at fluid inlet 25,across impeller portion 2 located toward the radial periphery of theassembly, across return vane 5, down and up through diffuser housing 3and out the exhaust 4 at a higher pressure than at the fluid inlet 25.Flow is from LEFT to RIGHT through the pump. Conversely, in the case ofa turbine which lowers the overall pressure of the liquid LNG, flow isfrom RIGHT to LEFT through the turbine.

The centrifugal pump comprises a rotatable shaft 1 rotating a pluralityof impellers 2 with fluid leaving the impeller to be diffused in thediffuser 3 and then passed to exhaust lines 4 which surround a submergedmotor housing 10. Fluid enters the impeller eye from a return-vane 5enclosed in a diffuser housing 6 all of which are encompassed in a pumphousing 7. The preceding impeller hub leakage is contained with anannular mechanical hub seal 8 which consists of a labyrinth and a smoothseal arrangement. The impeller eye is sealed with a mechanical shroudseal 9 using a shroud labyrinth or smooth seal arrangement. The impelleris circumferentially locked to the rotatable shaft with a key 14 andaxially a locknut 12. Behind the highest pressure impeller is the axialthrust equalizing mechanism consisting of a high pressure chamber 150and mechanical seal 100 through which pressurized fluid passes to thelow pressure chamber 204 and thrust plate 200. After passing throughthis low pressure chamber an axial variable orifice gap 203 is traverseby the thrust equalizing liquid and passes into the thrust plate pocket16 from where it exits the pocket through motor housing holes 19 orthrough the roller bearing 17 or through the bearing liner cooling holes300. The roller bearing is axially limited in travel with a locknut 15and washer 22 and spacer 23. Liquid which passes through the rollerbearing or bearing linear then passes through a motor housing bushing 21before entering the submerged cryogenic motor or generator cavity 20.

The destabilizing cross-coupled stiffness is a large influence on theforces that arise in mechanical seals and if too large can lead toexcessive synchronous and subsynchronous vibrations in centrifugal pumpand turbines.

The deswirl mechanisms claim in this invention serve two purposes.Firstly they act to deswirl liquid at the inlet of the mechanical sealswhich make up part of the thrust equalizing mechanism. Secondly thedeswirl mechanisms at the inlet of the variable axial orifice gap, alsopart of the axial thrust equalizing mechanism, removes unwantedcircumferential liquid velocity to avoid flow separation pockets whichgives a more stable axial thrust equalization than conventional liquidcryogenic systems. Together the mechanical seals and variable axialorifice gap act in harmony to equalize the axial thrust on the rotatingshaft. The present invention provides means for achieving the desirableinlet swirl reduction at two key locations in the axial thrustequalizing mechanism of cryogenic pumps and turbines.

Referring to the drawing and FIG. 2 in particular, shown therein is across section zoom of the region near 100. High pressure axial thrustequalizing liquid leaves the main core flow and travels inward along theback of the impeller in the annular high pressure chamber. Heresubstantial swirl is imparted on the liquid. The deswirl mechanisms 101in the high pressure chamber reduce and preset the circumferentialrotation of the liquid before it enters the clearance seal between thestationary wear ring 102 and the rotating labyrinth 103 mounted on theimpeller 2. The deswirl mechanisms are either vanes, fins, or groovescut into the material of the motor housing 10. Each of these deswirlmechanisms may be radial or inclined at an angle shown as α. In the caseof vanes or fins they are fastened to the motor housing 10 with bolts104 and are set to give a preswirl in the range 45°<α<135° with α=90° aspredominant. Testing and computational fluid dynamics with regard torotordynamic stability and in particular the stiffness and damping inthe seal optimizes the angle setting. After the thrust equalizing liquidleaves the seal and has undergone a substantial pressure drop in entersthe low pressure balance chamber 204. Since at the seal outlet theliquid will again have circumferential swirl a set of deswirl mechanisms201 are installed on the thrust plate 200. This deswirl mechanisms canbe fins, vanes, grooves or a combination thereof. The deswirl vanes orfins can be pivoted and locked into place with the bolts 202. The thrustequalizing liquid is then directed radially towards the variable axialorifice gap 203 where due to the lack of swirl it is more stable andavoids separation pockets. This serves for a more stable and improvedthrust equalizing mechanism. The impeller 2 is permitted to move axiallyin the range of 500 μm-3000 μm so that the axial orifice gap 203 isvariable. If an axial thrust is not equalized such that the axialorifice gap 203 begins to close the pressure in low pressure balancechamber 204 rises since the flow is restricted and there is lesspressure drop across the seal 102 and 103 from the high pressurechamber. This causes an increase in the axial opening force on the backof the impeller which counteracts and equalizes the closing axial thrustimbalance. If the axial thrust is not equalized in the reversedsituation such that the axial orifice gap 203 begins to open thepressure in low pressure balance chamber 204 decreases since the flow isless restricted and there is more pressure drop across the seal 102 and103 from the high pressure chamber. This causes a decrease in theopening force on the back of the impeller which counteracts andequalizes the opening axial thrust imbalance. After the axial thrustequalizing liquid leaves the axial orifice gap 203 it moves to towardthe roller bear 17 and the thrust plate pocket 16.

Referring to the drawing and FIG. 3 in particular, shown therein is across section zoom of the region near 100. High pressure axial thrustequalizing liquid leaves the main core flow and travels inward along theback of the impeller in the annular high pressure chamber. Heresubstantial swirl is imparted on the liquid. A plurality of holes at alarger radius 301 and smaller radius 302 are drilled into the motorhousing 10 where liquid by passes the wear ring gap inlet. The pluralityof holes are located about the circumference and radially staggered. Theholes and bypass liquid enter the clearance gap shortly downstream ofthe stationary wear ring 102 inlet and before the labyrinth rotatingwear ring 103. The holes are radially oriented so that liquid in theholes is of zero preswirl. After the axial thrust equalizing liquidleaves the seal and has undergone a substantial pressure drop it entersthe low pressure balance chamber 204 where it operates in harmony withthe variable axial orifice gap 203 as in the previous paragraphsdescribed manner.

Referring to the drawing and FIG. 4 in particular, shown therein is across section zoom of the region near 100. High pressure axial thrustequalizing liquid leaves the main core flow and travels inward along theback of the impeller in the annular high pressure chamber. Heresubstantial swirl is imparted on the liquid. The surface of the rotatingwear ring 103 is made smooth as mounted on the impeller 2. The surfaceof the stationary wear ring 102 is made up of a plurality of ridgesarranged in a diamond like pattern 400. These ridges 401 can be 1 mm to5 mm tall. They serve to brake the liquid swirl in the seal gap. Thediamond pattern is fixed annular type mounted inside the stationary wearring 102 which in turn is mounted into the motor housing 10. After theaxial thrust equalizing liquid leaves the seal and has undergone asubstantial pressure drop in enters the low pressure balance chamber 204where it operates in harmony with the variable axial orifice gap 203 asin the previous paragraphs described manner.

Referring to the drawing and FIG. 5 in particular, shown therein is across section zoom of the region near 300 for the situation after thethrust equalizing liquid leaves the low pressure chamber and variableorifice gap and enters the thrust plate chamber 16 of the thrust plate200. The liquid is blocked from entering the roller bearing 17 by abearing seal 502 that keeps low temperature lubricant 503 encapsulatedin the roller bearing. The roller bearing 17 is permitted to moveaxially approximately 500 μm-3000 μm in total to give the impeller axialtravel and vary the axial gap 203 depending on the axial thrust to beequalized. The roller bearing 17 is locked onto the shaft with thelocknut 15 washer 22 and spacer 23. The thrust equalizing liquid passesaround the roller bearing either passing though the bearing liner 501cooling slots 504 or the motor housing holes 19. If the liquid passesthrough the bearing liner cooling slots it then travels to the back ofthe roller bearing where it passes through the motor housing bushing 21and then to cool the submerged motor or generator. In this manner theroller bearing is completely lubricated with the low temperaturelubricant while the cryogenic liquid cools the bearing and lubricant.

Referring to the drawing and FIG. 6 in particular, shown therein is aaxial section zoom of the bearing liner 501. A plurality of axial groovecooling slots 504 and lands on the bearing liner 501 are used to passcooling liquid past the bearing. During start-up the cryogenic liquidmay be sufficiently cold to freeze the roller bearing lubricant. Abearing heater 505 is then need at start-up until the lubricant reachesnear −60° C. A temperature sensor 506 on the opposite side of the heateris applied to verify the start-up permission. The heater is applied fora few minutes before start-up of the pump or turbine. Afterward the heatfrom rotation in the roller bearing will keep the bearing lubricant warmand the heater can be shut-off.

Referring to the drawing and FIG. 7 in particular, shown therein is across section zoom of the region near 300 for the situation after thethrust equalizing liquid leaves the low pressure chamber and variableorifice gap and enters the thrust plate chamber 16 of the thrust plate200. The liquid is freely permitted to enter the roller bearing 17 whereit cools the bearing along the ball 705, inner race 704, and outer race702 as contact is made during rotation. The bearing cage material 703 isimpregnated with a dry lubricant that wipes and partially lubricates thebearing. The entire roller bearing 17 is permitted to move axiallyapproximately 500 μm-3000 μm in the bearing liner 701 to give theimpeller axial travel and vary the axial orifice gap 203 depending onthe axial thrust to be equalized. The roller bearing 17 is locked ontothe shaft with the locknut 15 washer 22 and spacer 23. The axial thrustequalizing liquid passes either through the roller bearing 17 or throughthe motor housing holes 19. If the liquid passes through the rollerbearing it then travels to the back of the roller bearing where itpasses through the motor housing bushing 21 and then to cool thesubmerged motor or generator. In this manner the roller bearing iscompletely cooled by flushing with low temperature thrust equalizingliquid.

Referring to the drawing and FIG. 8 in particular, shown therein is across section zoom of the impeller 2 and impeller eye seal region 9. Theimpeller shroud clearance leakage liquid passes through the mechanicallabyrinth seal with a plurality of teeth to the impeller eye. Normallyit enters the gap between the stationary smooth wear ring 801 embeddedin the diffuser housing 6 and the impeller eye wear ring 802 withsubstantial circumferential swirl. This swirl increases thedestabilizing cross-coupled stiffness. To eliminate this effect thecross-coupled stiffness is reduced using seal inlet deswirl mechanisms803. These are fins, vanes, or grooves. This operates to seal theimpeller and stabilize the rotordynamics in conjunction with the axialthrust equalizing mechanism.

Referring to the drawing and FIG. 9 in particular, shown therein is across section zoom of the impeller eye stationary seal wear rings 902and the smooth impeller eye wear ring 903, the operation of which isdescribed in the previous paragraph. On the stationary wear ring adiamond like surface pattern 901 like that shown previously in FIG. 4deswirls liquid in the clearance gap and reduces the cross-coupledstiffness. This seal operates to seal the impeller and stabilize therotordynamics in conjunction with the axial thrust equalizing mechanism.

Referring to the drawing and FIG. 10 in particular, shown therein is across section zoom of the impeller 2 and hub wear ring of region 8. Theimpeller hub clearance leakage liquid passes through the mechanicallabyrinth seal with a plurality of teeth to the impeller hub. Normallyit enters the gap between the stationary smooth wear ring 951 embeddedin the return-vane 5 and the impeller hub wear ring 952 with substantialcircumferential swirl. This swirl increases in the destabilizingcross-coupled stiffness. To eliminate this the cross-coupled stiffnessis reduced using inlet deswirl mechanisms 953. These are fins, vanes, orgrooves. This functions to seal the impeller hub clearance and stabilizethe rotordynamics in conjunction with the axial thrust equalizingmechanism.

Referring to the drawing and FIG. 11 in particular, shown therein is across section zoom of the return-vane stationary seal wear ring 975 andthe smooth impeller hub wear ring 977, the operation of which isdescribed in the previous paragraph. On the stationary wear ring adiamond like surface pattern 976 like that shown previously in FIG. 4 ismade to deswirl liquid in the clearance gap and reduce the cross-coupledstiffness. This seal operates to seal the impeller and stabilize therotordynamics in conjunction with the axial thrust equalizing mechanism.

The foregoing description is intended to illustrate the presentinvention. Those of ordinary skill will be able to envisage certainadditions, deletions or modifications to the described embodiments whichdo not depart from the spirit or scope of the invention as defined bythe claims herein.

Unless defined otherwise, all technical and scientific terms used hereinhave the same meaning as commonly understood by one of ordinary skill inthe art to which the present invention belongs. Although any methods andmaterials similar or equivalent to those described can be used in thepractice or testing of the present invention, the preferred methods andmaterials are now described. All publications and patent documentsreferenced in the present invention are incorporated herein byreference.

While the principles of the invention have been made clear inillustrative embodiments, there will be immediately obvious to thoseskilled in the art many modifications of structure, arrangement,proportions, the elements, materials, and components used in thepractice of the invention, and otherwise, which are particularly adaptedto specific environments and operative requirements without departingfrom those principles. The appended claims are intended to cover andembrace any and all such modifications, with the limits only of the truepurview, spirit and scope of the invention.

1. A pump with thrust equalizing mechanism for liquid cryogenic materials capable of operating at cryogenic liquid temperatures, the pump comprising, in part, a housing, a low pressure annular chamber in said housing to contain a low pressure liquid, a high pressure annular chamber in said housing to contain high pressure liquid, a rotating shaft concentric in said housing, said rotatable shaft constituting a rotatable element with a plurality of pump impellers mounted and rotating on a shaft connected to a submerged electric motor or generator, a liquid flow driven through a mechanical seal by a pressure difference from said high pressure chamber to said low pressure chamber, a first deswirl mechanism located inside said high pressure chamber arranged upstream of said mechanical seal to preset the preswirl of the liquid thrust equalizing flow that enters said seal to a predetermined predominantly radial inward direction, said first deswirl mechanism having largest radius inlet exposed to said high pressure chamber inlet, said first deswirl mechanism having outlet exposed to said mechanical seal inlet, such that said first deswirl mechanism deswirls said liquid thrust balancing flow swirl which was imparted by said rotating impeller to provide said seal with a preset inlet liquid flow swirl which may be zero in radial inward direction only, said mechanical seal exits liquid to said low pressure chamber, and a second deswirl mechanism positioned concentric with the rotatable shaft, said second deswirl mechanism arranged in a radial orientation upstream of an variable axial clearance to permit impeller rotation about said shaft center.
 2. The pump of claim 1, wherein the first deswirl mechanism comprises a plurality of vanes arranged about the circumference along the said rotatable shaft center, the plurality of vanes lying oriented in predetermined flow directions relative the location of the rotatable shaft center.
 3. The pump of claim 2, wherein the plurality of vanes are pivotable and can be locked into place in a predetermined position.
 4. The pump of claim 3, further comprising a controller and associated actuator, wherein the associated actuator can be used to control the direction of the plurality of pivotable vanes.
 5. The pump of claim 1, wherein the first deswirl mechanism comprises a plurality of fins arranged about the circumference along the said rotatable shaft center, the plurality of fins lying oriented in predetermined flow directions relative the location of the rotatable shaft center.
 6. The pump of claim 5, wherein the plurality of fins are pivotable and can be locked into place in a predetermined position.
 7. The pump of claim 6, further comprising a controller and associated actuator, wherein the associated actuator can be used to control the direction of the plurality of pivotable fins.
 8. The pump of claim 1, wherein the first deswirl mechanism comprises a plurality of grooves arranged about the circumference along the said rotatable shaft center, the plurality of grooves lying oriented in predetermined flow directions relative the location of the rotatable shaft center.
 9. The pump of claim 1, further comprising a primary plurality of liquid flow bypass passageway holes exiting said high pressure chamber, each of said liquid flow bypass passageway holes extending downstream to said seal with injection of deswirled liquid into the seal near said seal inlet at some intermediate pressure between that in said low high chamber and said low pressure chamber, said primary plurality of bypass holes having a predetermined radius.
 10. The pump of claim 9, further comprising a second plurality of liquid flow bypass passageway holes exiting said high pressure chamber, each of said liquid flow bypass passageway holes extending downstream to said seal with injection of deswirled liquid into the seal near said seal inlet at some intermediate pressure between that in said low high chamber and said low pressure chamber, said second plurality of bypass holes having a second predetermined radius.
 11. The pump of claim 1, wherein the mechanical seal is an annular mechanical seal to achieve pressure drop from said high pressure chamber to said low pressure chamber across said seal with rotating and stationary portions which is dependant on liquid flow rate through said seal, said seal rotating portion is a rotating labyrinth annulus positioned concentric with said rotatable shaft mounted on the highest pressure impeller stage, said labyrinth annulus consists of a plurality of circumferential grooved teeth with land and valley lengths, said seal stationary portion is smooth, distance between the rotating and stationary seal is the wear ring clearance wherein said liquid pressure drop results.
 12. The pump of claim 1, wherein the mechanical seal is an annular mechanical seal to achieve pressure drop from said high pressure chamber to said low pressure chamber across said mechanical seal with rotating and stationary portions which is dependant on liquid flow rate through said seal, said seal rotating portion is a smooth annulus positioned concentric with said rotatable shaft mounted on the highest pressure impeller stage, said seal stationary portion is a diamond surface pattern to act as a circumferential liquid flow deswirl mechanism, the distance between the rotating and diamond surface pattern stationary seal is the wear ring clearance wherein said liquid pressure drop results.
 13. The pump of claim 9, wherein the mechanical seal is an annular mechanical seal to achieve pressure drop from said high pressure chamber to said low pressure chamber across said seal with rotating and stationary portions which is dependant on liquid flow rate through said seal, said seal rotating portion is a rotating labyrinth annulus positioned concentric with said rotatable shaft mounted on the highest pressure impeller stage, said labyrinth annulus consists of a plurality of circumferential grooved teeth with land and valley lengths, said seal stationary portion is smooth, distance between the rotating and stationary seal is the wear ring clearance wherein said liquid pressure drop results.
 14. The pump of claim 10, wherein the mechanical seal is an annular mechanical seal to achieve pressure drop from said high pressure chamber to said low pressure chamber across said seal with rotating and stationary portions which is dependant on liquid flow rate through said seal, said seal rotating portion is a rotating labyrinth annulus positioned concentric with said rotatable shaft mounted on the highest pressure impeller stage, said labyrinth annulus consists of a plurality of circumferential grooved teeth with land and valley lengths, said seal stationary portion is smooth, distance between the rotating and stationary seal is the wear ring clearance wherein said liquid pressure drop results.
 15. The pump of claim 11 further comprising a second deswirl mechanism downstream of said liquid pressure drop apparatus comprising a plurality of fins to preset and adjust rotational swirl of said thrust equalizing liquid which exits said upstream seal and enters said low pressure chamber, itself upstream of a variable axial orifice gap.
 16. The pump of claim 12 further comprising a second deswirl mechanism downstream of said liquid pressure drop apparatus comprising a plurality of fins to preset and adjust rotational swirl of said thrust equalizing liquid which exits said upstream seal and enters said low pressure chamber, itself upstream of a variable axial orifice gap.
 17. The pump of claim 11 further comprising a second deswirl mechanism downstream of said liquid pressure drop apparatus comprising a plurality of vanes to preset and adjust rotational swirl of said thrust equalizing liquid which exits said upstream seal and enters said low pressure chamber, itself upstream of a variable axial orifice gap.
 18. The pump of claim 12 further comprising a second deswirl mechanism downstream of said liquid pressure drop apparatus comprising a plurality of vanes to preset and adjust rotational swirl of said thrust equalizing liquid which exits said upstream seal and enters said low pressure chamber, itself upstream of a variable axial orifice gap.
 19. The pump of claim 11 further comprising a second deswirl mechanism downstream of said liquid pressure drop apparatus comprising a plurality of grooves to preset and adjust rotational swirl of said thrust equalizing liquid which exits said upstream seal and enters said low pressure chamber, itself upstream of a variable axial orifice gap.
 20. The pump of claim 12 further comprising a second deswirl mechanism downstream of said liquid pressure drop apparatus comprising a plurality of grooves to preset and adjust rotational swirl of said thrust equalizing liquid which exits said upstream seal and enters said low pressure chamber, itself upstream of a variable axial orifice gap.
 21. The pump of claim 20 wherein the liquid cryogenic apparatus further comprises an axial gap of variable axial gap size capable of axial movement acting as a variable orifice to constitute a variable liquid flow restriction based on the axial location of said rotating shaft, the axial gap comprising a rotating and stationary smooth surface with a variable axial orifice gap, said rotating surface coupled to the neighboring highest pressure impeller, said rotating surface able to move axially acting as the variable side of a variable orifice, said rotating surface making up one side of a radially orientated axial gap, said stationary surface as the other side of a radially orientated variable axial orifice gap.
 22. The pump of claim 21 further comprising a variable pressure chamber controlled with said variable axial orifice gap, the variable pressure chamber further comprising the second deswirl mechanism.
 23. The pump of claim 22 further comprising a liquid cryogenic roller bearing assembly functioning in tandem and conjunction with said first and second liquid deswirl mechanisms and said variable axial orifice gap, the roller bearing assembly comprising an unsealed roller bearing cooled with thrust equalizing liquid flow flushing through, said unsealed bearing lubricated with a dry impregnated lubricant bearing cage, said unsealed bearing accepting a fraction of the thrust equalizing liquid from said variable orifice mechanism with remaining unwanted liquid flow bypassing, said unsealed bearing located concentric with outer race inside a bearing liner with a small radial clearance of between about 10 μm and about 60 μm to permit said unsealed bearing to move axially with said variable orifice gap, said bearing liner is fixed in a stationary housing.
 24. The pump of claim 22 further comprising a liquid cryogenic roller bearing assembly functioning in tandem and conjunction with said first and second deswirl mechanisms and said variable axial orifice gap, the roller bearing assembly comprising a sealed roller bearing packed permanently with low temperature lubricant, said sealed bearing located with outer race concentric inside a bearing liner with a small radial clearance of between about 10 μm and about 60 μm to permit said sealed bearing to move axially with said variable orifice mechanism, said sealed bearing accepting no through liquid flow, said bearing liner fixed in a stationary housing, said bearing liner further having a plurality of grooved axial slots about the circumference to pass a fraction of liquid flow from said variable orifice mechanism for cooling, said sealed roller bearing apparatus with a bearing start-up heater located near said bearing liner, bearing temperature sensor mounted circumferentially about 180 degrees or more or less from said bearing heater, the pump further comprising a start-up delay control system whereby said bearing heater is activated to preheat said roller bearing lubricant to a predetermined temperature before start-up is permitted.
 25. The pump of claim 22, wherein the cryogenic liquid mechanical seal assembly comprises a plurality of impeller eye wear rings functioning in conjunction and harmony with the first and second deswirl mechanisms as part of the thrust equalizing mechanism, the thrust equalizing mechanism comprising a rotating labyrinth seal with a plurality of circumferential grooved teeth on the rotating impeller wear ring, the thrust equalizing mechanism further comprising a stationary smooth surface which together with said rotating wear ring forms a radial clearance gap to seal impeller shroud leakage fluid, the thrust equalizing mechanism further comprising a plurality of tertiary deswirl mechanisms upstream of the seal.
 26. The pump of claim 22, wherein the cryogenic liquid mechanical seal assembly comprises a plurality of impeller eye wear rings functioning in conjunction and harmony with the first and second deswirl mechanisms as part of the thrust equalizing mechanism in cryogenic liquids comprising a rotating annular smooth seal on the rotating impeller or rotating impeller wear ring, a stationary diamond pattern mesh surface which together with said rotating wear ring forms a radial clearance gap to seal impeller shroud leakage liquid and deswirl liquid in the clearance gap, the thrust equalizing mechanism further comprising a plurality of deswirl mechanisms upstream of the seal.
 27. The pump of claim 1 further comprising a cryogenic liquid mechanical seal assembly on the plurality of impeller interstage bushings and wear rings functioning in conjunction and harmony with first and second deswirl mechanisms as part of the thrust equalizing mechanism, the mechanical seal assembly comprising a stationary smooth surface annular wear ring mounted in a fixed housing, a rotating labyrinth seal with a plurality of circumferential grooved teeth on a rotating impeller or rotating impeller annular wear ring which together with said stationary wear ring forms a radial clearance gap to seal interstage return liquid, the thrust equalizing mechanism further comprising a plurality of tertiary deswirl upstream of the seal.
 28. A pump of claim 1 further comprising a cryogenic liquid mechanical seal assembly on the plurality of impeller interstage bushings and wear rings functioning in conjunction and harmony with first and second deswirl mechanisms as part of the thrust equalizing mechanism, the mechanical seal assembly comprising a stationary diamond pattern mesh surface on an annular surface seal, a rotating smooth surface on the impeller or impeller wear ring which together with said stationary wear ring forms a radial clearance gap to seal interstage return liquid, a plurality of tertiary deswirl mechanisms upstream of the seal. 